ISSN: [Mukherjee * et al., 6(9): September, 2017] Impact Factor: 4.116

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1 IC Value:.00 IJESRT INTERNATIONAL JOURNAL OF ENGINEERING SCIENCES & RESEARCH TECHNOLOGY DESIGN AND ANALYSIS OF POWER TRAIN SYSTEM OF HEAVY TRUCK ENGINE Sabyasachi Mukherjee* & Puspendu Chandra Assistant Professor, Department of Mechanical Engineering, Regent Education and Research Foundation, North Parganas, Barrackpore, Kolkata, West Bengal, India DOI: /zenodo ABSTRACT A powertrain is a system of mechanical parts in a vehicle that takes the power, or output, of the engine and, through specific gear ratios, slows it and transmits it as torque. Through the driveshaft, the engine s torque is transmitted to the wheels of the vehicle, which, when applied to road, moves the body. Simply put, a powertrain is made up of a transmission system and a driveshaft.the mechanism that transmits the power developed by the engine of the automobile to the driving wheels is called the transmission system (or Power train).it provides a varied leverage between the engine and the drive wheels. It also provides the connection and disconnection of engine with rest of power train without shock and smoothly.the average person is most familiar with the powertrain of their car, which creates energy in the engine, which is transferred to the transmission. Main objective of this project is to design a powertrain system for a truck engine by manual calculation and computer aided designing. KEYWORDS: Powertrain system, Transmission, Torque, Computer aided designing, Design Calculations. I. INTRODUCTION A powertrain is a system of mechanical parts in a vehicle that first transmits the power developed by the engine of the automobile to the driving wheels. The average person is most familiar with the powertrain of their car, which creates energy in the engine, which is transferred to the transmission. The transmission then takes the power, or output, of the engine and, through specific gear ratios, slows it and transmits it as torque. Through the driveshaft, the engine s torque is transmitted to the wheels of the car, which, when applied to road, moves the car. Simply put, a powertrain is made up of a transmission and a driveshaft. [] TATA Prima LX truck engine has been taken for powertrain design purpose. [7] Volume of the engine : 588 cc Power : 500 rpm (0r) rpm Torque : rpm Gears : 6 speed gear box( 5F + 1R) Clutch: Single plate dry friction clutch. Main parts of powertrain system are: Clutch Gear box Propeller shaft Universal joint Differential Layout of a powertrain system is shown below in Fig. 1. It consists of clutch, gearbox, propeller shaft, universal joint and differential unit. [51]

2 IC Value:.00 Fig. 1 Layout diagram of Powertrain System II. DESIGN CALCULATIONS Design Calculation for Single Plate Clutch Clutch is designed based on two theories: Uniform pressure theory Uniform wear theory Uniform Pressure Theory Torque can be determined by using the formula below T = μ W R T= Torquetransmitted by engine. μ =Co-efficient of friction of material. Fa =Axial force = π p ri(r o -r i ). R=Effective mean radius= / (r o - r i / r o r i ) Pr = intensity of pressure = p r i = p r o (p wil be maximum where r is minimum= r i ) Centre distance layout of a single plate clutch is given in fig. below Dimensions Fig Centre Distance of Clutch Plate By using design parameters we get the dimensions of single plate clutch The outer diameter of the clutch plate do=0 mm The inner diameter of the clutch plate is di=00 mm. N / mm [5]

3 IC Value:.00 Intensity of pressure = 0.1 Co-efficient of friction, Asbestos material=0.5 Calculation of Axial force Axial force can be determined by using the formula below F ππp(r π Torque transmitted by Clutch a i o F a 9. 7N r ) The formula of torque is given below T WR T 5. 5Nm Mean Radius of Friction Surface, R = 0.107m i Dimensions of the springs Total Load on the springs, 1.5F a N Since there are 6 springs Therefore maximum load on each spring is, We know that Wahls Stress factors, C=6 C K 1.55 C C We also know that Maximum shear stress induced in the wire is 0Mpa We shall take the Standard wire of size SWG8 having diameter, \Mean diameter of the spring, Let us assume that the spring has active turns (n=) Therefore compression of the spring, mm Assuming squared and ground ends, total Number of turns ' n n 6 We know that free length of the spring W s 196. N 6 K 8W S C d d 119.6d d. 99mm d. 51mm D C. d mm 8W s G.C d N G 810 mm [5] n

4 IC Value:.00 L f mm ' n ' L f n d mm And pitch of the Coils, Design of Six Speed Gear Box I assumed the minimum and maximum speed of a six speed gear box is 500 r.p.m and 500 r.p.m respectively. Selection of Spindle Speeds n 6; N r. p. m min 500 N 500 max r. p. m n1 N N max min We Know that, 61 We find =1.0 is not a standard ratio, satisfies the requirement. Therefore the sipngle speeds from R0 series, skipping two speeds are given as, Structural Formula: (1). () ,710, 1000,150,1800, 500 r.p.m Fig Ray diagram [5]

5 IC Value:.00 Kinematic Arrangement for 6-Speed gear box Groups Z 1 + Z = Z + Z ; Z 5 + Z 6 = Z 7 + Z 8 = Z 9 + Z 10 Design of Spur gears C5 steel material is selected (PSG DB, Page 8.5) Fig Kinematic Arrangement Design stress, [σ b ] = 150 kgf/cm, [σ c ] = 5000 kgf/cm E= kgf/cm Ψ=0.ψ m =10 We assume Z 1 =0, kk d =1. Centre distance between output shaft and input shaft Nominal twisting moment transmitted, [M t ] = 970 kw n (PSG DB, Page 1)a (i + 1) [[ (0.7) [σ c ] ] E [M t ] iψ ] 1. = 178 kgf-cm I=1000/500= a 16cm or 160 mm [M t ] Module m 1.6 [ ] 0.75cm y [σ b ]ψ m z From PSG DB Page 8., module=8 mm y=0.8 Calculation of number of teeth Z 1 = a m(i + 1) = 160 8( + 1) = 1. ; Z = i Z 1 = Z 1 + Z = Z + Z = 0 ; Z + iz =0; Z Z = i = 1.5 Calculation of gear Diameter Z = 18 ; Z = d 1 = mz 1 = 10mm ; d = 16mm ; d = 1mm; d = 176mm [55]

6 IC Value:.00 Centre Distance between Intermediate shaft and Input shaft [M t ] = 970 kw n Z 5 = a m(i+1) = 5 ; Z 6 = 5 1. = kgf-cm Z 7 + Z 8 = 70 ; Z 7 + iz 8 = 70 i = Z 8 Z 7 = 1.8 a 8.5 cm or 85mm Z 7 = 9 ; Z 8 = 1 Z 9 + Z 10 = 70 ; Z 9 + iz 10 = 70 ; Z 9 = 8; Z 10 = i = Z 10 Z 9 = 1. d 5 = 00mm; d 6 =60mm; d 7 = mm; d 8 = 8mm; d 9 =mm; d 10 = 6mm Calculation of Revised Centre distance Centre distance b/w output and intermediate shaft, Centre distance b/w intermediate and input shaft, Calculation of face width d1 d a1 160mm d5 d6 a 80mm b m 80 b= 80 mm Calculation of Length of shaft Length of the Shaft, L= (0+10+b+0+7b+7b+10+0)mm, L=100+18b=150mm Design of shafts Design of output shaft Fn L 1) To find Maximum bending moment (M); M the lowest speed is 500 r.p.m is obtained by meshing gears 1 and T 1 = P 60 = = 86.6 N m πn π 500 F Fn = Normal load on gear = t N cos cos0 Maximum B.M, M N mm To find the Equivalent torque (Teq) The formula is given below T eq M Diameter of the Spindle T From PSG book table 9.5 d s 16T eq / N 0 mm Rounded off values of the diameter, using R0 series is 15mm 1/ T eq mm 1mm 6 N mm [56]

7 IC Value:.00 Selection of Bearing According to the spindle diameter, Deep Groove Ball Bearing SKF606 8 has been selected. From PSG data book Series 60, Page.1 Design of Universal Coupling In designing a universal coupling, the shaft diameter and the pin diameter is obtained and discussed below, d = diameter of shaft dp = Diameter of pin τ1= Allowable shear stress for the material of the shaft and pin respectively. Torque transmitted by the shafts T d 16 Since the pin is in double shear, therefore the torque transmitted T d p 1 d Assuming the torque transmitted by the engine is The allowable shear stresses for the shaft and pin may be taken as 60 MPa and 0 MPa respectively. Diameter of Shaft Diameter of Pin T Design of Propeller Shaft 6 N mm T d d 16 T d p 1 d d p d d mm 106mm 6 d p 5.95mm 5mm The assumptions are shaft rotates at a constant speed about its longitudinal axis. The shaft has a uniform, circular cross section. The shaft is perfectly balanced, i.e., at every cross section, the mass center coincides with the geometric center. All damping and nonlinear effects are excluded. Taking a simple drive shaft or propeller shaft is designed using following assumed data, Maximum Torque (T) = 500 N-m Length of the shaft L = 150 mm Inclination angle ( ) = Deg Density = 7600 Kg m Yield Stress in shear = 70 Mpa Rotational speed (N) = 6500 rpm Young s Modulus = 810 Mpa To start with, take the diameter of driving shaft is 50 mm The assumed diameter is then will be used as input for the torsional shear stress [57]

8 IC Value:.00 Calculation of polar moment of inertia (I) d 50 I mm Where, d = diameter of driving shaft Calculate the Maximum torsional shear stress, S max Td Smax 1. Mpa I We know Yield strength for driveshaft material is 70 Mpa (C5) So, we can conclude the propeller shaft is safe for the transmitted torque in the shaft. Calculate Maximum static deflection of the drive shaft 5. m. g.cos. L 8. E. I COS mm Where, m = mass of the propeller shaft = kg g = acceleration due to gravity = 9.81 m/s = Inclination angle = deg L = length of the propeller shaft = 150 mm E = Young s modulus = 8 10 Mpa Calculate the Critical speed of the Shaft ( N c 0 g N c rpm. ) Therefore it is seen that speed of the drive shaft is 6500 r.p.m, which is lower than the calculated critical speed of shaft(6696rpm). Design is safe. Design of Differential By using PSG Data book, the following formulas are listed below Formula Used No of Teeth = Z Module, m t = 5mm Cone Distance, R = 0.5m t (z 1 + z ) Outer Pitch Diameter, d = m t.z Tip Diameter d a1 = m t (z+ cos) Face width, b = 10.m t Pitch angle, tan = i Tip angle, a = + a Root angle, f = - a Addendum, h a = m t Dedendum, h f = 1.88m t Addendum angle, tan a = (m t f 0 )/R Dedendum angle, tan f = (m t (f 0 +c))/r [58]

9 IC Value:.00 Height factor = 1 Clearance = 0. mm Pinion Gear Calculation Pinion Gear It is the small Gear in the differential unit No of Teeth = 16 teeth Module, m t = 5mm Cone Distance, R = 16.91mm Outer Pitch Diameter, d = 80mm Tip Diameter d a1 = 89.86mm Face width, b = 0mm Pitch angle, = 18. Tip angle, a = 0.7 Root angle, f = Addendum, h a = 5mm Dedendum, h f = 9.mm Addendum angle, a =.6 Dedendum angle, f =.715 Height factor = 1 Clearance = 0. mm Ring Gear Calculation Ring Gear It is largest size gear in the differential unit No of Teeth = 8 teeth Module, m t = 5mm Cone Distance, R = 16.91mm Outer Pitch Diameter, d = 0mm Tip Diameter d a1 =.16mm Face width, b = 0mm Pitch angle, = Tip angle, a = 7.8 Root angle, f = Addendum, h a = 5mm Dedendum, h f = 9.mm Addendum angle, a =.6 Dedendum angle, f =.715 Height factor = 1 Clearance = 0. mm Identical Bevel Gear (miter gear) Calculation Miter Gear - A type of bevel gear used in pairs with intersecting shafts at 90 angles. Both the driving gear and driven gear in a miter gear pair have the same diameter, same number of teeth, and a mechanical advantage of 1. No of Teeth = 18 teeth Module, m t = 5mm Cone Distance, R = 6.69mm Outer Pitch Diameter, d = 90mm Tip Diameter d a1 = mm Face width, b = 0mm Pitch angle, = 5 [59]

IC Value:.00 Tip angle, a = 9.9 Root angle, f = 9.61 Addendum, h a = 5mm Dedendum, h f = 9.mm Addendum angle, a =.9 Dedendum angle, f = 5.86 Height factor = 1 Clearance = 0. mm III.

5 below Modeling of Propeller Shaft Fig 5 Universal Coupling Assembled diagram Modeling of propeller shaft is done using CATIA V5 software and assembled with Universal coupling.

10 IC Value:.00 Tip angle, a = 9.9 Root angle, f = 9.61 Addendum, h a = 5mm Dedendum, h f = 9.mm Addendum angle, a =.9 Dedendum angle, f = 5.86 Height factor = 1 Clearance = 0. mm III. MODELING Modeling of Universal Coupling Modeling of Universal coupling is done by using CATIA V5 software and shown in fig.5 below Modeling of Propeller Shaft Fig 5 Universal Coupling Assembled diagram Modeling of propeller shaft is done using CATIA V5 software and assembled with Universal coupling. The assembled model is shown in fig. 6 below Fig 6. CATIA Model of Propeller Shaft IV. ANALYSIS Analysis of Propeller Shaft Using ANSYS 1.1 Analysis of the Propeller Shaft is done using ANSYS 1.1 software. First the model is meshed properly and then it is analyzed. The meshed model is shown below in fig. 7 [60]

IC Value:.00 Fig 7. Meshed Model Next 70 N/m load is applied and total deformation is observed.maximum and minimum deformation is also determined. The total deformation diagram is given below in fig.

11 IC Value:.00 Fig 7. Meshed Model Next 70 N/m load is applied and total deformation is observed.maximum and minimum deformation is also determined. The total deformation diagram is given below in fig. 8 Fig 8. Total deformation in Propeller shaft Below in fig.9 the equivalent stress model (von-misses) is given, where we can determine the maximum and minimum stress induced for the loading condition. Fig 9. Equivalent Stress model [61]

12 IC Value:.00 V. RESULTS Results of Single Plate Clutch Axial force Torque transmitted by Clutch F a 9. 7N T 5. 5Nm Total Load on the springs is N Standard wire of size SWG8 having diameter, d. 51mm Compression of the spring, mm Free length of the spring, ' L f n d mm Pitch of the Coils5.0 mm Results of Six Speed Gear box Centre distance b/w output and intermediate shaft is 160 mm Centre distance b/w intermediate and input shaft is 80 mm Face width, b=80 mm Length of shaft, L= 150 mm Diameter of spindle, d=15mm According to the spindle diameter, Deep Groove Ball Bearing SKF606 8 has been selected Results of Universal Coupling Diameter of shaft, D=106 mm Diameter of pin, d=5mm Results of Propeller Shaft Polar moment of Inertia, d 50 I mm Maximum Torsional Shear Stress, Td S s max 1. Mpa I Maximum static deflection of the drive shaft Critical speed of the Shaft ( N c )= 6696 r.m.p Results of Differential = mm All the results of Differential unit are given in the table no. 1 below [6]

13 IC Value:.00 Table No.1: Results of Differential Unit Components Pinion Gear Ring Gear Milter Gear No. of teeth Module (mm) Cone distance(mm) Outer pitch diameter (mm) Tip diameter(mm) Face width (mm) Pitch angle (degree) Tip angle (degree) Root angle(degree) Addendum (mm) Dedendum (mm) Addendum angle(degree) Dedendum (degree) angle Height factor Clearance VI. CONCLUSION The results of this report have illustrated the entire design methodology into the powertrain system assembly. The efforts taken to account for all the necessary design and analysis considerations have provided a solid starting point into the fundamentals of powertrain system design. The opportunity to participate in the global collaboration project has provided tremendous insight into the nature of cooperation essential to the successful completion of a multi-faceted project. The design procedure includes the following design calculations Design of Clutch Design of 6 speed gear box Design of Universal coupling Design of Propeller shaft Design of Differential unit Using CATIA V5 modelling software, the propeller shaft design is done.analysis of Total Deformation and Equivalent Stress is also done by using ANSYS 1.1 VII. REFERENCES [1] KalaikathirAchchagam, Faculty of mechanical engineering, PSG Design Data Book, Revised Edition December 01 [] Khurmi R.S and Gupta J.K, Machine Design, Eurasia Publishing House, First Multi Colour Edition 005 [6]

14 IC Value:.00 [] Khurmi R.S and Gupta J.K, Theory of Machines, S.CHAND Publication, Second Edition 00 [] Jayakumar.V, Design of Transmission Systems, S.CHAND Publication, Third Edition 008 [5] Khanna O.P, Production Technology, Anuradha Publication, Revised Third Edition 005 [6] Prabhu T.J, Design of Transmission Elements, Devi Xerox, Fifth Edition 000 [7] Tata Prima Truck Specifications, [8] Wikipedia, CITE AN ARTICLE Mukherjee, S., & Chandra, P. (017). DESIGN AND ANALYSIS OF POWER TRAIN SYSTEM OF HEAVY TRUCK ENGINE. INTERNATIONAL JOURNAL OF ENGINEERING SCIENCES & RESEARCH TECHNOLOGY, 6(9), [6]

Design and Vibrational Analysis of Flexible Coupling (Pin-type)

Design and Vibrational Analysis of Flexible Coupling (Pin-type) 1 S.BASKARAN, ARUN.S 1 Assistant professor Department of Mechanical Engineering, KSR Institute for Engineering and Technology, Tiruchengode,